Compressor with motor cooling

ABSTRACT

A compressor includes a compression mechanism, a shaft, a motor, and a cooling medium delivery structure. The motor includes a rotor mounted on the shaft and a stator disposed radially outwardly of the rotor to form a gap between the rotor and the stator. The cooling medium delivery structure includes inlet and outlet conduits located to supply and discharge a cooling medium to and from the motor. The shaft has an external shape different than an internal shape of the rotor to form at least one axial passageway between the shaft and the rotor. The cooling medium is supplied through the gap and the at least one axial passageway to cool the rotor.

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a continuation application of U.S. patentapplication Ser. No. 15/072,975, allowed, filed on Mar. 17, 2016. Theentire disclosure of U.S. patent application Ser. No. 15/072,975 ishereby incorporated herein by reference.

BACKGROUND Field of the Invention

The present invention generally relates to a compressor. Morespecifically, the present invention relates to a compressor with motorcooling.

Background Information

A chiller system is a refrigerating machine or apparatus that removesheat from a medium. Commonly a liquid such as water is used as themedium and the chiller system operates in a vapor-compressionrefrigeration cycle. This liquid can then be circulated through a heatexchanger to cool air or equipment as required. As a necessarybyproduct, refrigeration creates waste heat that must be exhausted toambient or, for greater efficiency, recovered for heating purposes. Aconventional chiller system often utilizes a centrifugal compressor,which is often referred to as a turbo compressor. Thus, such chillersystems can be referred to as turbo chillers. Alternatively, other typesof compressors, e.g. a screw compressor, can be utilized.

In a conventional (turbo) chiller, refrigerant is compressed in thecentrifugal compressor and sent to a heat exchanger in which heatexchange occurs between the refrigerant and a heat exchange medium(liquid). This heat exchanger is referred to as a condenser because therefrigerant condenses in this heat exchanger. As a result, heat istransferred to the medium (liquid) so that the medium is heated.Refrigerant exiting the condenser is expanded by an expansion valve andsent to another heat exchanger in which heat exchange occurs between therefrigerant and a heat exchange medium (liquid). This heat exchanger isreferred to as an evaporator because refrigerant is heated (evaporated)in this heat exchanger. As a result, heat is transferred from the medium(liquid) to the refrigerant, and the liquid is chilled. The refrigerantfrom the evaporator is then returned to the centrifugal compressor andthe cycle is repeated. The liquid utilized is often water.

A conventional centrifugal compressor basically includes a casing, aninlet guide vane, an impeller, a diffuser, a motor, various sensors anda controller. Refrigerant flows in order through the inlet guide vane,the impeller and the diffuser. Thus, the inlet guide vane is coupled toa gas intake port of the centrifugal compressor while the diffuser iscoupled to a gas outlet port of the impeller. The inlet guide vanecontrols the flow rate of refrigerant gas into the impeller. Theimpeller increases the velocity of refrigerant gas. The diffuser worksto transform the velocity of refrigerant gas (dynamic pressure), givenby the impeller, into (static) pressure. The motor rotates the impeller.The controller controls the motor, the inlet guide vane and theexpansion valve. In this manner, the refrigerant is compressed in aconventional centrifugal compressor. A conventional centrifugalcompressor may have one or two stages. A motor drives the one or moreimpellers.

The motor in the conventional centrifugal compressor may need to becooled. The general method of motor cooling is by used refrigerant ofthe chiller system. See for example U.S. Pat. No. 3,805,547, U.S. Pat.No. 3,645,112, and Japanese publication No. JPH01-138946.

SUMMARY

One example of a relatively common refrigerant used in a centrifugalchiller system is R134a. The conventional motor cooling techniques workrelatively well when this refrigerant is used in a conventional chillersystem. See FIGS. 25-26. However, it has been discovered that when lowpressure refrigerant (LPR), for example R1233zd, is used in acentrifugal chiller system the conventional motor cooling techniques maynot be sufficient. See FIGS. 25 and 27.

Therefore an object of the present invention is to provide a centrifugalcompressor for a chiller that adequately cools the motor even when LPRis used such as R1233zd.

It has also been discovered that rotor temperature gets higher thanstator temperature in a conventional centrifugal chiller system usingconventional motor cooling techniques when a low pressure refrigerant(LPR) such as R1233zd is used. See FIG. 27. The temperature of the rotorand/or the stator can also become higher than desired.

Therefore another object of the present invention is to provide acentrifugal compressor for a chiller that adequately cools the rotorand/or stator of the motor even when LPR is used such as R1233zd.

It has further been discovered that an amount of motor cooling dependson refrigerant flow rate, and that too high of refrigerant flow rate inthe convention centrifugal compressor can result in drag on the motor.

Therefore another object of the present invention is to provide acentrifugal compressor for a chiller in which adequate refrigerant flowrate is provided without causing drag on the motor.

It has further been discovered that a refrigerant flow rate, and thus,an amount of cooling depends on a pressure difference between higher andlower sides. Pressure difference of R134a is higher than LPR such asR1233zd. However it has been further discovered that the cross sectionalarea of the flow path is also a factor in the refrigerant flow rate.

Therefore another object of the present invention is to provide acentrifugal compressor for a chiller in which an adequate crosssectional area of a flow path and/or a pressure difference is providedto facilitate adequate refrigerant flow and cooling even when a lowpressure refrigerant (LPR) such as R1233zd is used.

It has further been discovered that a low global warming potential (GWP)refrigerant such as R1234ze or R1234yf can also have one or more of theabove challenges.

Therefore, yet another object of the present invention is to provide acentrifugal compressor for a chiller in which low global warmingpotential (GWP) refrigerant such as R1234ze or R1234yf can be used toadequately cool the motor in accordance with one or more of the aboveobjects.

One or more of the foregoing objects can basically be achieved byproviding a compressor including a compression mechanism, a shaftrotatable about a rotation axis and attached to the compressionmechanism to rotate a part of the compression mechanism, a motor and acooling medium delivery structure. The motor is arranged to rotate theshaft. The motor includes a rotor mounted on the shaft and a statordisposed radially outwardly of the rotor to form a gap between the rotorand the stator. The cooling medium delivery structure includes an inletconduit located to supply a cooling medium to the motor and an outletconduit located to discharge the cooling medium from the motor. Theshaft has an external shape different than an internal shape of therotor to form at least one axial passageway between the shaft and therotor along an axial length of the shaft at least as long as an axiallength of the rotor. The inlet conduit is located to supply the coolingmedium through the gap and the at least one axial passageway to cool therotor. The outlet conduit is located to discharge the cooling mediumfrom the gap and the at least one axial passageway.

These and other objects, features, aspects and advantages of the presentinvention will become apparent to those skilled in the art from thefollowing detailed description, which, taken in conjunction with theannexed drawings, discloses preferred embodiments.

BRIEF DESCRIPTION OF THE DRAWINGS

Referring now to the attached drawings which form a part of thisoriginal disclosure:

FIG. 1 is a schematic diagram illustrating a single stage chiller systemhaving a centrifugal compressor in accordance with an embodiment of thepresent invention;

FIG. 2 is a schematic diagram illustrating a two stage chiller system(with an economizer) having a centrifugal compressor in accordance withan embodiment of the present invention;

FIG. 3 is partial schematic diagram illustrating a first option ofstator and rotor cooling flow paths applicable to the chiller systems ofFIGS. 1 and 2, with the second stage shown in a phantom circle toillustrate that the second stage is only present in FIG. 2;

FIG. 4 is partial schematic diagram illustrating a second option ofstator and rotor cooling flow paths applicable to the chiller system ofFIG. 2;

FIG. 5 is partial schematic diagram illustrating a third option ofstator and rotor cooling flow paths applicable to the chiller systems ofFIGS. 1 and 2, with the second stage shown in a phantom circle toillustrate that the second stage is only present in FIG. 2;

FIG. 6 is partial schematic diagram illustrating a fourth option ofstator and rotor cooling flow paths applicable to the chiller systems ofFIGS. 1 and 2, with the second stage shown in a phantom circle toillustrate that the second stage is only present in FIG. 2;

FIG. 7 is partial schematic diagram illustrating a fifth option ofstator and rotor cooling flow paths applicable to the chiller systems ofFIGS. 1 and 2, with the second stage shown in a phantom circle toillustrate that the second stage is only present in FIG. 2;

FIG. 8 is partial schematic diagram illustrating a combination of thefirst and third options of stator and rotor cooling flow pathsapplicable to the chiller systems of FIGS. 1 and 2, with the secondstage shown in a phantom circle to illustrate that the second stage isonly present in FIG. 2;

FIG. 9 is partial schematic diagram illustrating a combination of thefirst and fourth options of stator and rotor cooling flow pathsapplicable to the chiller systems of FIGS. 1 and 2, with the secondstage shown in a phantom circle to illustrate that the second stage isonly present in FIG. 2;

FIG. 10 is partial schematic diagram illustrating a combination of thefirst and fifth options of stator and rotor cooling flow pathsapplicable to the chiller systems of FIGS. 1 and 2, with the secondstage shown in a phantom circle to illustrate that the second stage isonly present in FIG. 2;

FIG. 11 is partial schematic diagram illustrating a combination of thesecond and third options of stator and rotor cooling flow pathsapplicable to the chiller system of FIG. 2;

FIG. 12 is partial schematic diagram illustrating a combination of thesecond and fourth options of stator and rotor cooling flow pathsapplicable to the chiller system of FIG. 2;

FIG. 13 is partial schematic diagram illustrating a combination of thesecond and fifth options of stator and rotor cooling flow pathsapplicable to the chiller system of FIG. 2;

FIG. 14 is a perspective view of the centrifugal compressor of thechiller system illustrated in FIG. 2, with portions broken away andshown in cross-section for the purpose of illustration;

FIG. 15A is a simplified partial longitudinal cross-sectional view ofthe motor of the compressors illustrated in FIGS. 1-14 illustrating afirst parallel directional flow of rotor cooling;

FIG. 15B is a simplified partial longitudinal cross-sectional view ofthe motor of the compressors illustrated in FIGS. 1-14 illustrating asecond parallel directional flow of rotor cooling;

FIG. 16A is a simplified partial longitudinal cross-sectional view ofthe motor of the compressors illustrated in FIGS. 1-14 illustrating afirst series directional flow of rotor cooling;

FIG. 16B is a simplified partial longitudinal cross-sectional view ofthe motor of the compressors illustrated in FIGS. 1-14 illustrating asecond series directional flow of rotor cooling;

FIG. 17 is a schematic longitudinal cross-sectional view of theimpellers, motor and magnetic bearings of the centrifugal compressorillustrated in FIGS. 1-16, with cooling medium flow omitted for the sakeof simplicity;

FIG. 18 is an enlarged perspective view of the motor shaft of the motorof the compressors illustrated in FIGS. 1-17;

FIG. 19 is a longitudinal elevational view of the motor shaftillustrated in FIG. 18;

FIG. 20 is an end elevational view of the motor shaft illustrated inFIGS. 18-19;

FIG. 21 is a transverse cross-sectional view of the motor shaftillustrated in FIGS. 18-20 as viewed along section line 21-21 of FIG.19;

FIG. 22 is a transverse cross-sectional view of the motor shaftillustrated in FIGS. 18-20 as viewed along section line 22-22 of FIG.19;

FIG. 23 is a partial transverse cross-section view of the motor of thecompressor illustrated In FIGS. 15A-15B, as seen along section line23-23 of FIG. 15A, illustrating a negative angle of the grooves in theshaft relative to the rotation direction;

FIG. 24 is a partial transverse cross-section view of the motor of thecompressor illustrated In FIGS. 15A-15B, as seen along section line23-23 of FIG. 15A, illustrating a positive angle of the grooves in theshaft relative to the rotation direction;

FIG. 25 is partial cross-sectional view of the centrifugal compressorhaving a conventional motor;

FIG. 26 is a chart illustrating stator and rotor temperatures in theconventional compressor of FIG. 25, when R134a is the refrigerant;

FIG. 27 is a chart illustrating stator and rotor temperatures in theconventional compressor of FIG. 25, when R1233zd is the refrigerant; and

FIG. 28 is a chart illustrating stator and rotor temperatures in theconventional compressor of FIG. 25, when R134a is the refrigerant.

DETAILED DESCRIPTION OF EMBODIMENT(S)

Selected embodiments will now be explained with reference to thedrawings. It will be apparent to those skilled in the art from thisdisclosure that the following descriptions of the embodiments areprovided for illustration only and not for the purpose of limiting theinvention as defined by the appended claims and their equivalents.

Referring initially to FIGS. 1 and 2, chiller systems 10 and 10′ havingcentrifugal compressors 22 and 22′ in accordance with an embodiment ofthe present invention are illustrated. The centrifugal compressor 22 ofFIG. 1 is a single stage compressor, and thus, the chiller system 10 ofFIG. 1 is a single stage chiller system. The centrifugal compressor 22′of FIG. 2 is a two stage compressor, and thus, the chiller system 10′ ofFIG. 2 is a two stage chiller system. The two stage chiller system ofFIG. 2 also includes an economizer. FIGS. 1 and 2 merely illustrate twoexamples of chiller systems in which centrifugal compressors 22 and 22′in accordance with the present invention can be used.

Referring now briefly to FIGS. 3-13 numerous options for attaching thecentrifugal compressors 22 and 22′ in the chiller systems 10 and 10′ inorder to provide motor cooling flow in accordance with the presentinvention are illustrated. FIGS. 1 and 2 do not illustrate the motorcooling flows shown in FIGS. 3-13 because of the numerous options shownin FIGS. 3-13, if included in FIGS. 1-2, could make FIGS. 1-2 confusing.However, it will be apparent to those skilled in the art from thisdisclosure that the options of FIGS. 3-13 can be incorporated in thechiller systems 10 and 10′ illustrated in FIGS. 1 and 2 as indicatedabove in the Brief Descriptions of the Drawings. In addition it will beapparent to those skilled in the art from this disclosure that theeconomizer of the chiller system 10′ can be eliminated when not used formotor cooling flow in FIGS. 3-13.

The chiller systems 10 and 10′ are conventional, except for thecentrifugal compressors 22 and 22′ and the manner in which the coolingflows are supplied to the centrifugal compressors 22 and 22′. Thereforethe chiller systems 10 and 10′ will not be discussed and/or illustratedin detail herein except as related to the centrifugal compressors 22 and22′ and the manner in which the cooling flows are supplied to thecentrifugal compressors 22 and 22′. However, it will be apparent tothose skilled in the art that the conventional parts of the chillersystems 10 and 10′ can be constructed in variety of ways withoutdeparting the scope of the present invention. In the illustratedembodiments, the chiller systems 10 and 10′ are preferably waterchillers that utilize cooling water and chiller water in a conventionalmanner.

The centrifugal compressors 22 and 22′ are identical to each other,except the centrifugal compressor 22′ is a two stage compressor. Thus,it will be apparent to those skilled in the art from this disclosurethat the singe stage compressor 22 is identical to the centrifugalcompressor 22′, except for the removal of parts. Therefore, the twostage compressor 22′ includes all the parts of the single stagecompressor 22, but also includes additional parts. Accordingly, it willbe apparent to those skilled in the art from this disclosure that thedescriptions and illustrations of the two stage compressor 22′ alsoapply to the single stage compressor 22, except for parts relating tothe second stage of compression and modifications related to the secondstage of compression (e.g., the housing shape, shaft end shape, etc.).In view of these points, and for the sake of brevity, only the two stagecompressor 22′ will be explained and/or illustrated in detail herein.The compressor 22′ will be explained in more detail below.

Referring again to FIGS. 1-2, the components of the chiller systems 10and 10′ will now briefly be explained. The chiller system 10 basicallyincludes a chiller controller 20, the compressor 22, a condenser 24, anexpansion valve or orifice 27, and an evaporator 28 connected togetherin series to form a loop refrigeration cycle. The chiller system 10′includes a chiller controller 20, the centrifugal compressor 22′, acondenser 24, an expansion valve or orifice 25, an economizer 26, anexpansion valve or orifice 27, and an evaporator 28 connected togetherin series to form a loop refrigeration cycle. In either case, varioussensors (not shown) are disposed throughout the circuits of the chillersystems 10 and 10′ to control the chiller systems 10 and 10′ in aconventional manner.

Referring now to FIGS. 1-17, mainly FIGS. 14-17, the compressor 22′ willnow be explained in more detail. The compressor 22′ is a two-stagecentrifugal compressor in the illustrated embodiment. Thus, thecompressor 22′ illustrated herein includes two impellers. However, thecompressor 22′ include three or more impellers (not shown) or may be asingle stage compressor as shown in FIG. 1. The two-stage centrifugalcompressor 22′ of the illustrated embodiment is conventional except thatthe compressor 22′ includes motor cooling paths connected to thecompressor 22′ as shown in one of FIGS. 3-13, and cooling refrigerantsupplied within the compressor 22′ as shown in FIGS. 15A-15B. Of course,it will be apparent to those skilled in the art from this disclosurethat the cooling paths of FIGS. 16A-16B could also be used withoutdeparting from the scope of the present invention. The motor coolingwill be explained in more detail below.

Thus, the centrifugal compressor 22′ includes a first stage impeller 34a and a second stage impeller 34 b. The centrifugal compressor 22′further includes a first stage inlet guide vane 32 a, a firstdiffuser/volute 36 a, a second stage inlet guide vane 32 b, a seconddiffuser/volute 36 b, a compressor motor 38, and a magnetic bearingassembly 40 as well as various conventional sensors (only some shown).While magnetic bearings are described herein, it will be apparent tothose skilled in the art from this disclosure that other types and formsof compressor bearings maybe used with this invention. A casing 30covers the other parts of the centrifugal compressor 22′. The casing 30includes an inlet portion 31 a and an outlet portion 33 a for the firststage of the compressor 22′. The casing 30 also includes an inletportion 31 b and an outlet portion 33 b for the second stage of thecompressor 22′.

The chiller controller 20 receives signals from the various sensors andcontrols the inlet guide vanes 32 a and 32 b, the compressor motor 38,and the magnetic bearing assembly 40 in a conventional manner, asexplained in more detail below. Refrigerant flows in order through thefirst stage inlet guide vane 32 a, the first stage impeller 34 a, thesecond stage inlet guide vane 32 b, and the second stage impeller 34 b.The inlet guide vanes 32 a and 32 b control the flow rate of refrigerantgas into the impellers 34 a and 34 b, respectively, in a conventionalmanner. The impellers 34 a and 34 b increase the velocity of refrigerantgas, generally without changing pressure. The motor speed determines theamount of increase of the velocity of refrigerant gas. Thediffusers/volutes 36 a and 36 b increase the refrigerant pressure. Thediffusers/volutes 36 a and 36 b are non-movably fixed relative to thecasing 30. The compressor motor 38 rotates the impellers 34 a and 34 bvia a shaft 42. The magnetic bearing assembly 40 magnetically supportsthe shaft 42. Alternatively, the bearing system may include a rollerelement, a hydrodynamic bearing, a hydrostatic bearing, and/or amagnetic bearing, or any combination of these. In this manner, therefrigerant is compressed in the centrifugal compressor 22′.

In operation of the chiller system 10, the first stage impeller 34 a andthe second stage impeller 34 b of the compressor 22′ are rotated, andthe refrigerant of low pressure in the chiller system 10 is sucked bythe first stage impeller 34 a. The flow rate of the refrigerant isadjusted by the inlet guide vane 32 a. The refrigerant sucked by thefirst stage impeller 34 a is compressed to intermediate pressure, therefrigerant pressure is increased by the first diffuser/volute 36 a, andthe refrigerant is then introduced to the second stage impeller 34 b.The flow rate of the refrigerant is adjusted by the inlet guide vane 32b. The second stage impeller 34 b compresses the refrigerant ofintermediate pressure to high pressure, and the refrigerant pressure isincreased by the second diffuser/volute 36 b. The high pressure gasrefrigerant is then discharged to the chiller system 10.

Referring to FIGS. 14-17, the magnetic bearing assembly 40 isconventional, and thus, will not be discussed and/or illustrated indetail herein, except as related to the present invention. Rather, itwill be apparent to those skilled in the art that any suitable magneticbearing can be used without departing from the present invention. Themagnetic bearing assembly 40 preferably includes a first radial magneticbearing 44, a second radial magnetic bearing 46 and an axial (thrust)magnetic bearing 48. In any case, at least one radial magnetic bearing44 or 46 rotatably supports the shaft 42. The thrust magnetic bearing 48supports the shaft 42 along a rotational axis X by acting on a thrustdisk 45. The thrust magnetic bearing 48 includes the thrust disk 45which is attached to the shaft 42.

The thrust disk 45 extends radially from the shaft 42 in a directionperpendicular to the rotational axis X, and is fixed relative to theshaft 42. A position of the shaft 42 along rotational axis X (an axialposition) is controlled by an axial position of the thrust disk 45. Thefirst and second radial magnetic bearings 44 and 46 are disposed onopposite axial ends of the compressor motor 38. Various sensors detectradial and axial positions of the shaft 42 relative to the magneticbearings 44, 46 and 48, and send signals to the chiller controller 20 ina conventional manner. The chiller controller 20 then controls theelectrical current sent to the magnetic bearings 44, 46 and 48 in aconventional manner to maintain the shaft 42 in the correct position.The magnetic bearing assembly 40 is preferably a combination of activemagnetic bearings 44, 46, and 48, which utilizes gap sensors 54, 56 and58 to monitor shaft position and send signals indicative of shaftposition to the chiller controller 20. Thus, each of the magneticbearings 44, 46 and 48 are preferably active magnetic bearings.

Referring now to FIGS. 14-24, the motor 38 in accordance with thepresent invention will now be explained in more detail. The motor 38includes a stator 60 and a rotor 62. The stator 60 is fixed to aninterior surface of the casing 30. On the other hand, the rotor 62 isfixed to the shaft 42. The stator 60 and the rotor 62 are conventional.Thus, when electricity is sent to the stator 60, the rotor 62 is causedto rotate. Since the rotor is fixed to the shaft 42, the shaft 42 isalso caused to rotate, and thus, the impellers 34 a and 34 b are alsocause to rotate. A gap G is formed between the stator 60 and the rotor62. The gap G extends circumferentially completely around the rotor 62and axially along the lengths of the stator 60 and rotor 62. Coolingfluid is supplied to the outside of the stator 60. In addition, coolingfluid is supplied to an axial end of the motor 38 to cool the rotor 62by passing axially through the gap G. Cooling of the stator 60 and therotor 62 will be explained in more detail below.

Referring to FIGS. 18-24, the shaft 42 will now be explained in moredetail. It should be noted that FIG. 17 is a simplified view, and thus,does not illustrate the portions of the shaft 42. As mentioned above,the rotor 62 is mounted on the shaft 42. The shaft 42 includes a firstradial magnetic bearing portion 64, a second radial magnetic bearingportion 66, a third axial magnetic bearing support portion 68, anenlarged portion 70 and a rotor support portion 72. In addition,impeller support portions 74 a and 74 b are disposed at opposite ends ofthe shaft 42 and have the impellers 34 a and 34 b fixedly attachedthereto.

The first radial magnetic bearing portion 64 is axially disposed betweenthe rotor support portion 72 and impeller support portion 34 a. Thefirst radial magnetic bearing portion 64 is magnetically radiallysupported by the first radial magnetic bearing 44 in a conventionalmanner. The third axial magnetic bearing support portion 68 is axiallydisposed between the enlarged portion 70 and second magnetic bearingportion 66. The third axial magnetic bearing support portion 68 has thethrust disk 45 fixedly mounted thereon in a conventional manner (notshown in FIGS. 18-24). The thrust disk 45 is axially magneticallysupported by the axial magnetic bearing 48 in a conventional manner. Thesecond magnetic bearing portion 66 is axially disposed between the thirdaxial magnetic bearing support portion 68 and the second impellersupport portion 74 b. The second radial magnetic bearing portion 66 ismagnetically radially supported by the second radial magnetic bearing 46in a conventional manner.

The rotor support portion 72 is axially disposed between the firstmagnetic bearing portion 64 and the enlarged portion 70. The enlargedportion 70 is axially disposed between the rotor support portion 72 andthe third axial magnetic bearing support portion 68. A plurality ofgrooves 80 are formed in the outside surface of portions of the enlargedportion 70 and the rotor support portion 72. Due to the presence of thegrooves 80, the shaft 42 has an external shape different from aninternal shape of the rotor 62 to form a plurality of axial passageways.Due to the enlarged portion 70 being larger than the rotor supportportion 72, the rotor 62 can be slid onto the rotor support portion 72until the rotor 62 contacts the enlarged portion 70. See FIGS. 14-16.However the grooves 80 have lengths longer than the rotor 62 and extendalong part of the enlarged portion 70 and the rotor support portion 72.In addition, the grooves 80 have depths larger than the difference inradial height between the enlarged portion 70 and the rotor supportportion 72, as best understood from FIGS. 18 and 21. Thus, cooling fluidcan pass axially through the grooves 80, as explained in more detailbelow. The cooling fluid can pass in parallel as shown in FIG. 15A fromleft to right, or in parallel as shown in FIG. 15B from right to left.Alternatively, the cooling medium can pass in series *e.g., a counterflow) as shown in FIG. 16A to/from the left, or in series as shown inFIG. 16B to/from the right. The flows of FIGS. 16A and 16B can beparticularly useful when there is a large pressure difference betweenthe supply side and the return side.

Referring still to FIGS. 18-24 the grooves 80 of the shaft 42 will nowbe explained in more detail. In the illustrated embodiment, the shaft 42has six grooves 80 equally circumferentially spaced from each other.Thus, the external shape of the shaft 42 includes an annular section (ofeach of the enlarged portion 70 and rotor support portion 72) and aplurality of grooves 80 extending radially inwardly from the annularsections. In addition, in the illustrated embodiment, the grooves 80 areidentical to each other. Each groove 80 includes a first sidewall 82, asecond sidewall 84 circumferentially spaced from the first sidewall 82and a trough wall 86 connecting radially inner ends of the first andsecond sidewalls 82 and 84. The first sidewall 82 of each groove 80 issubstantially parallel to the second sidewall 84 of the groove 80 asviewed in axial cross section. In addition, each groove 80 has acenterline C equally spaced from the first and second sidewalls 82 and84 as viewed in axial cross section, and the centerline C of each groove80 is inclined relative to a radial direction of the shaft 42 as bestunderstood from FIGS. 23 and 24.

In view of the above configuration, the external shape of the shaft 42is different than the internal shape of the rotor 62 to form a pluralityof axial passageways between the shaft 42 and the rotor 62 along theaxial length of the shaft 42 at least as long as the axial length of therotor 62. In any event, the shaft 42 has an external shape differentthan an internal shape of the rotor 62 to form at least one axialpassageway between the shaft 42 and the rotor 62 along an axial lengthof the shaft 42 at least as long as an axial length of the rotor 62.When discussed the axial length of the rotor 62 here it is intended torefer to an axial length of the portion of the rotor 62 attached to theshaft 42. A total cross sectional area of the at least one axialpassageway is larger than a total cross sectional area of the gap G asviewed in axial cross section. Preferably, the total cross sectionalarea of the at least one axial passageway is approximately double thetotal cross sectional area of the gap G as viewed in axial crosssection.

The shaft 42 rotates in a rotation direction R during operation of thecentrifugal compressor 22 or 22′, and each centerline C may be inclinedso that a radially inner end is disposed circumferentially further inthe positive rotational direction than a radially outer end of thecenterline C as shown in FIG. 23. This illustrates a negative angle ofthe grooves 80. Alternatively, the shaft 42 rotates in a rotationdirection R during operation of the centrifugal compressor 22 or 22′,and each centerline C may be inclined so that a radially outer end isdisposed circumferentially further in the positive rotational directionthan a radially inner end of the centerline C as shown in FIG. 24. Thisillustrates a positive angle of the grooves.

The groove configurations of the illustrated embodiment are merelyexamples. However, it will be apparent to those skilled in the art fromthis disclosure that the exact groove configuration may be calculatedbased on fluid simulation, the characteristics of the groove. However,it is preferable that there is an angle with respect to the rotationaldirection. The direction of such an angle can be determined as follows.When designed with an emphasis on cooling, a “negative angle” withrespect to the rotational direction as shown in FIG. 23 may be mostuseful, while when designed with an emphasis on friction loss, a“positive angle” with respect to the rotational direction, because therotational resistance (friction loss) of the shaft increases in the caseof the “positive angle” with respect to the rotational direction asshown in FIG. 24 may be the most useful.

In either case, it is preferable that the total area of the shaftgrooves 80 is approximately double the passage area of the air gap G.Thus, a total cross sectional area of the plurality of grooves 80 islarger than a total cross sectional area of the gap G as viewed in axialcross section. Preferably, the total cross sectional area of theplurality of grooves 80 is approximately double the total crosssectional area of the gap G as viewed in axial cross section. In theillustrated embodiment a ratio of the grooves 80 area to the gap G areais 0.63 to 0.37. However, the optimum groove area as well as the grooveangle is preferably determined in accordance with whether designed withan emphasis on cooling or on decreasing friction loss. In discussing thecross-section areas herein it should be noted that the gap G istypically very small so that its size is enlarged herein for the sake ofillustration.

Referring again to FIGS. 3-13, the options for cooling medium deliveryto the motor 38 will be explained in more detail. In all of FIGS. 3-13there is provided a stator supply SS, a stator return SR, at least onerotor supply RS and a rotor return RR. While only illustrated as linesin these figures, these lines represent conventional conduits/piping asbest understood from FIGS. 15A-15B and 16A-16B. The RR and RS lines inFIGS. 15A-15B may be combined only at the solid lines. In other words asingle rotor supply line RS and a single Rotor return line RR can beprovided in FIGS. 15A-15B or two parallel lines may be provided. Ineither case, FIGS. 15A-15B illustrated parallel flow through the gap Gand the grooves 80. In FIGS. 16A-16B, series flow through the grooves 80and the gap G is illustrated to/from opposite ends of the motor 38.Thus, only a single rotor supply line RS and a single rotor return lineRR is used. It will be apparent to those skilled in the art from thisdisclosure that at least any of the flows of FIGS. 15A, 15B, 16A, 16Bcan be used with the flows of FIGS. 3-13 without departing from thescope of the present invention.

In FIGS. 3-13, the stator supply lines SS and stator return lines SR arethe same for all of FIGS. 3-13. Each stator supply line SS includes twosolenoid valves SOV sandwiching a dryer filter DF therebetween. Eachstator return line SR includes a solenoid valve SOV. In addition, therotor return line RR for each of FIGS. 3-13 is also the same. However,the rotor supply lines RS for FIGS. 3-13 are different. In FIGS. 3-13,some of the cooling medium delivery arrangements apply to the firstand/or second chiller systems 10 or 10′ in which case the second stageof the compressor 22′ is surrounded by hidden lines, to indicate that itis optional. In these cases, the rotor supply line RS is not impacted bythe presence or absence of the second stage of the compressor 22′

In FIG. 3 the rotor supply line RS delivers cooling fluid from theevaporator 28 to the motor 38. Thus, this delivery applies to the singlestage chiller 10 or the two stage chiller system 10′.

In FIG. 4 the rotor supply line RS delivers cooling fluid from theeconomizer 26 to the motor 38. Thus, this delivery applies to the twostage chiller system 10′.

In FIG. 5 the rotor supply line RS delivers cooling fluid from thecondenser 24 to the motor 38. In this option, the rotor supply line RSincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith an expansion valve EXV downstream. Thus, this delivery applies tothe single stage chiller 10 or the two stage chiller system 10′.

In FIG. 6 the rotor supply line RS delivers cooling fluid from thecondenser 24 to the motor 38. In this option, the rotor supply line RSincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith an Orifice O downstream. Thus, this delivery applies to the singlestage chiller 10 or the two stage chiller system 10′.

In FIG. 7 the rotor supply line RS delivers cooling fluid from thecondenser 24 to the motor 38. In this option, the rotor supply line RSincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith parallel mounted expansion valve EXV and orifice O downstream.Thus, this delivery applies to the single stage chiller 10 or the twostage chiller system 10′.

In FIG. 8 the rotor supply line RS delivers cooling fluid from thecondenser 24 and from the evaporator 28 to the motor 38 via a branchpoint. In this option, the rotor supply line RS from the condenserincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith an expansion valve EXV downstream. Thus, this delivery applies tothe single stage chiller 10 or the two stage chiller system 10′.

In FIG. 9 the rotor supply line RS delivers cooling fluid from thecondenser 24 and from the evaporator 28 to the motor 38 via a branchpoint. In this option, the rotor supply line RS from the condenserincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith an Orifice O downstream. Thus, this delivery applies to the singlestage chiller 10 or the two stage chiller system 10′.

In FIG. 10 the rotor supply line RS delivers cooling fluid from thecondenser 24 and from the evaporator 28 to the motor 38 via a branchpoint. In this option, the rotor supply line RS from the condenserincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith parallel mounted expansion valve EXV and orifice O downstream.Thus, this delivery applies to the single stage chiller 10 or the twostage chiller system 10′.

In FIG. 11 the rotor supply line RS delivers cooling fluid from thecondenser 24 and from the economizer 26 to the motor 38 via a branchpoint. In this option, the rotor supply line RS from the condenserincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith an expansion valve EXV downstream. Thus, this delivery applies tothe single stage chiller 10 or the two stage chiller system 10′.

In FIG. 12 the rotor supply line RS delivers cooling fluid from thecondenser 24 and from the economizer 26 to the motor 38 via a branchpoint. In this option, the rotor supply line RS from the condenserincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith an Orifice O downstream. Thus, this delivery applies to the twostage chiller system 10′.

In FIG. 13 the rotor supply line RS delivers cooling fluid from thecondenser 24 and from the economizer 26 to the motor 38 via a branchpoint. In this option, the rotor supply line RS from the condenserincludes solenoid valves SOV sandwiching a strainer ST therebetween, andwith parallel mounted expansion valve EXV and orifice O downstream.Thus, this delivery applies to the two stage chiller system 10′.

The controller 20 may control the valves and/or the orifice size may beset to delivery the correct amount of refrigerant. The stator supply SS,the stator return SR, the least one rotor supply RS and the rotor returnRR as well as the parts disposed therein form parts of a cooling mediumdelivery structure in accordance with the present invention. The coolingmedium delivery structure further includes an inlet conduit IC locatedto supply the cooling medium to a first axial end of the motor 38 and anoutlet conduit OC located to discharge the cooling medium from a secondaxial end of the motor 38, as best understood from FIGS. 3-16. The inletconduit IC is located to supply the cooling medium from the first axialend of the motor 38 through the gap G and the at least one axialpassageway (e.g., formed by the grooves 80) to the second axial end ofthe motor 38 to cool the rotor 62, and the outlet conduit OC is locatedto discharge the cooling medium supplied to the second axial end of themotor 38 from the gap G and the at least one axial passageway (e.g.,formed by the grooves 80). Of course, the first and second axial endscan be reversed as shown in FIGS. 15A-15B and 16A-16B. The conduits ICand OC can supply/discharge cooling medium to/from both the gap G andthe grooves 80, or additional conventional conduits (e.g., like thosepreviously used to deliver fluid to a gap) can be provided for the gapG.

In the illustrated embodiment, the cooling medium delivery structuredoes not include a pump. In addition, in the illustrated embodiment, atleast a portion of the inlet conduit (IC) is axially disposed closer toone of the first or third magnetic bearing elements 44 or 48 than thefirst axial end of the motor 38, depending on which direction thecooling medium flows. At least a portion of the outlet conduit OC isaxially disposed closer to one of the second or third magnetic bearingelements 46 or 48 than the second axial end of the motor 38, dependingon which direction the cooling medium flows. In the illustratedembodiment, the third axial thrust bearing element 48 is axiallydisposed between one of the first and second radial magnetic bearingelements 44 and 46 and one of the first and second axial ends of themotor 38, respectively.

Referring to FIGS. 1 and 2, the chiller controller 20 may includenumerous control sections programmed to control the conventional partsin a conventional manner. For example, a conventional magnetic bearingcontrol section, a conventional compressor variable frequency drive, aconventional compressor motor control section, a conventional inletguide vane control section, and a conventional expansion valve controlsection. These sections can be separate or combined sections.

In the illustrated embodiment, the control sections are sections of thechiller controller 20 programmed to execute the control of the partsdescribed herein. However, it will be apparent to those skilled in theart from this disclosure that the precise number, location and/orstructure of the control sections, portions and/or chiller controller 20can be changed without departing from the present invention so long asthe one or more controllers are programed to execute control of theparts of the chiller system 10 as explained herein.

The chiller controller 20 is conventional, and thus, includes at leastone microprocessor or CPU, an Input/output (I/O) interface, RandomAccess Memory (RAM), Read Only Memory (ROM), a storage device (eithertemporary or permanent) forming a computer readable medium programmed toexecute one or more control programs to control the chiller system 10.The chiller controller 20 may optionally include an input interface suchas a keypad to receive inputs from a user and a display device used todisplay various parameters to a user. The parts and programming areconventional, and thus, will not be discussed in detail herein, exceptas needed to understand the embodiment(s).

In terms of global environment protection, use of new low GWP (GlobalWarming Potential) refrigerants such like R1233zd, R1234ze areconsidered for chiller systems. One example of the low global warmingpotential refrigerant is low pressure refrigerant in which theevaporation pressure is equal to or less than the atmospheric pressure.For example, low pressure refrigerant R1233zd is a candidate forcentrifugal chiller applications because it is non-flammable, non-toxic,low cost, and has a high COP compared to other candidates such likeR1234ze, which are current major refrigerant R134a alternatives. In theillustrated embodiment, the cooling medium is refrigerant used in thechiller system 10 or 10′. Preferably the refrigerant is at least one ofa low pressure refrigerant (LPR) and a low global warming potential(GWP) refrigerant. More specifically, the low pressure refrigerant (LPR)may be R1233zd and/or the low global warming potential (GWP) refrigerantmay be R1234ze or R1234yf.

Referring now to FIGS. 26-28, a conventional motor using R134a isillustrated in FIG. 26, a conventional motor using R1233zd isillustrated in FIG. 27, and a motor in accordance with the presentinvention using R1233zd is illustrated in FIG. 28. As shown in theFigure, when cooling is performed in the same shape, sufficient coolingis performed, for example, with respect to the temperature limit of 60°C. in R134a as shown in the FIG. 26; however, the temperature limit isexceeded in R1233zd as shown in the FIG. 27. This is because R134a iscapable of supplying the refrigerant supply amount of, for example, 0.48kg/s, whereas R1233zd is only capable of supplying 0.18 kg/s(approximately one-third). The difference in pressure (high pressure−lowpressure) is used in supplying the refrigerant; thus, the absolutesupply amount decrease in R1233zd. Also, when a groove is absent, therotor cooling heat transfer part will be the outer surface of the rotoronly. This lack of heat transfer area influences the temperatureincrease in R1233zd. As shown in FIG. 28, when a groove is provided inthe shaft, the inner temperature of the motor 38 will be sufficientlycooled. By providing a groove in the shaft, the passage area inside therefrigerant increases and the refrigerant supply amount increases. Byproviding a groove in the shaft, the rotor is cooled from the outsideand from the inside of the rotor.

Regarding the distance from the magnetic bearing—normally, magneticbearings are disposed at three positions in total, that is, radialmagnetic bearings 1, 2 and thrust magnetic bearing 3. For cooling of themagnetic bearing, the following arrangement is preferred as shown inFIGS. 15A-15B and 16A-16B. Arrange the refrigerant supply port (e.g.,inlet conduit IC) and the exhaust port (e.g., outlet conduit OC) asclose as possible to the magnetic bearing. Arrange the shaft groove 80and the magnetic bearing such that the distance between them is also assmall as possible. As shown in the examples of FIGS. 15A-15B and16A-16B, the positions of the refrigerant supply/exhaust ports can beswitched in accordance with the rotor load or the load on each magneticbearing, and such switching can be performed during operation.

GENERAL INTERPRETATION OF TERMS

In understanding the scope of the present invention, the term“comprising” and its derivatives, as used herein, are intended to beopen ended terms that specify the presence of the stated features,elements, components, groups, integers, and/or steps, but do not excludethe presence of other unstated features, elements, components, groups,integers and/or steps. The foregoing also applies to words havingsimilar meanings such as the terms, “including”, “having” and theirderivatives. Also, the terms “part,” “section,” “portion,” “member” or“element” when used in the singular can have the dual meaning of asingle part or a plurality of parts.

The term “detect” as used herein to describe an operation or functioncarried out by a component, a section, a device or the like includes acomponent, a section, a device or the like that does not requirephysical detection, but rather includes determining, measuring,modeling, predicting or computing or the like to carry out the operationor function.

The term “configured” as used herein to describe a component, section orpart of a device includes hardware and/or software that is constructedand/or programmed to carry out the desired function.

The terms of degree such as “substantially”, “about” and “approximately”as used herein mean a reasonable amount of deviation of the modifiedterm such that the end result is not significantly changed.

While only selected embodiments have been chosen to illustrate thepresent invention, it will be apparent to those skilled in the art fromthis disclosure that various changes and modifications can be madeherein without departing from the scope of the invention as defined inthe appended claims. For example, the size, shape, location ororientation of the various components can be changed as needed and/ordesired. Components that are shown directly connected or contacting eachother can have intermediate structures disposed between them. Thefunctions of one element can be performed by two, and vice versa. Thestructures and functions of one embodiment can be adopted in anotherembodiment. It is not necessary for all advantages to be present in aparticular embodiment at the same time. Every feature which is uniquefrom the prior art, alone or in combination with other features, alsoshould be considered a separate description of further inventions by theapplicant, including the structural and/or functional concepts embodiedby such feature(s). Thus, the foregoing descriptions of the embodimentsaccording to the present invention are provided for illustration only,and not for the purpose of limiting the invention as defined by theappended claims and their equivalents.

What is claimed is:
 1. A compressor comprising: a compression mechanismthat compresses refrigerant when rotated; a shaft rotatable about arotation axis and attached to the compression mechanism to rotate a partof the compression mechanism; a motor arranged to rotate the shaft, themotor including a rotor mounted on the shaft and a stator disposedradially outwardly of the rotor to form a gap between the rotor and thestator; and a cooling medium delivery structure including an inletconduit located to supply a cooling medium to the motor and an outletconduit located to discharge the cooling medium from the motor, theshaft having an external shape different than an internal shape of therotor to form at least one axial passageway between the shaft and therotor along an axial length of the shaft at least as long as an axiallength of the rotor, and the inlet conduit being located to supply thecooling medium through the gap and the at least one axial passageway tocool the rotor, and the outlet conduit being located to discharge thecooling medium from the gap and the at least one axial passageway. 2.The compressor according to claim 1, wherein the external shape of theshaft is different than the internal shape of the rotor to form aplurality of axial passageways between the shaft and the rotor along theaxial length of the shaft at least as long as the axial length of therotor.
 3. The compressor according to claim 2, wherein the externalshape of the shaft includes an annular section and a plurality ofgrooves extending radially inwardly from the annular section.
 4. Thecompressor according to claim 3, wherein the grooves are substantiallyequally spaced from each other along a circumferential direction aboutthe annular section.
 5. The compressor according to claim 3, whereineach of the grooves includes a first sidewall, a second sidewallcircumferentially spaced from the first sidewall and a trough wallconnecting radially inner ends of the first and second sidewalls.
 6. Thecompressor according to claim 5, wherein the first sidewall of eachgroove is substantially parallel to the second sidewall of the groove asviewed in axial cross section.
 7. The compressor according to claim 5,wherein each groove has a centerline equally spaced from the first andsecond sidewalls as viewed in axial cross section, and the centerline ofeach groove is inclined relative to a radial direction of the shaft. 8.The compressor according to claim 7, wherein the shaft rotates in arotation direction during operation, and each centerline is inclined sothat a radially outer end is disposed circumferentially further in thepositive rotational direction than a radially inner end of thecenterline.
 9. The compressor according to claim 7, wherein the shaftrotates in a rotation direction during operation, and each centerline isinclined so that a radially inner end is disposed circumferentiallyfurther in the positive rotational direction than a radially outer endof the centerline.
 10. The compressor according to claim 5, wherein theshaft rotates in a rotation direction during operation, and at least oneof the first sidewall and the second sidewall of each groove is angledso that a radially outer end is disposed circumferentially further inthe positive rotational direction than a radially inner end.
 11. Thecompressor according to claim 5, wherein the shaft rotates in a rotationdirection during operation, and at least one of the first sidewall andthe second sidewall of each groove is angled so that a radially innerend is disposed circumferentially further in the positive rotationaldirection than a radially outer end.
 12. The compressor according toclaim 3, wherein a total cross sectional area of the plurality ofgrooves is larger than a total cross sectional area of the gap as viewedin axial cross section.
 13. The compressor according to claim 12,wherein the total cross sectional area of the plurality of grooves isapproximately double the total cross sectional area of the gap as viewedin axial cross section.
 14. The compressor according to claim 1, whereina total cross sectional area of the at least one axial passageway islarger than a total cross sectional area of the gap as viewed in axialcross section.
 15. The compressor according to claim 14, wherein thetotal cross sectional area of the at least one axial passageway isapproximately double the total cross sectional area of the gap as viewedin axial cross section.
 16. The compressor according to claim 1, whereinthe cooling medium is refrigerant used in compressor.
 17. The compressoraccording to claim 16, wherein the refrigerant is a low pressurerefrigerant (LPR).
 18. The compressor according to claim 17, wherein thelow pressure refrigerant (LPR) is R1233zd.
 19. The compressor accordingto claim 16, wherein the refrigerant is a low global warming potential(GWP) refrigerant.
 20. The compressor according to claim 19, wherein thelow global warming potential (GWP) refrigerant is R1234ze or R1234yf.21. The compressor according to claim 16, wherein the cooling mediumdelivery structure does not include a pump.
 22. The compressor accordingto claim 16, further comprising a magnetic bearing rotatably supportingthe shaft.
 23. The compressor according to claim 22, wherein themagnetic bearing includes a first radial magnetic bearing elementdisposed on a first axial end of the motor, a second radial magneticbearing element disposed on a second axial end of the motor, and a thirdaxial thrust bearing element disposed on one of the first and secondaxial ends of the motor.
 24. The compressor according to claim 23,wherein at least a portion of the inlet conduit is axially disposedcloser to one of the first or third magnetic bearing elements than thefirst axial end of the motor.
 25. The compressor according to claim 23,wherein at least a portion of the outlet conduit is axially disposedcloser to one of the second or third magnetic bearing elements than thesecond axial end of the motor.